Percussive unit for earth drilling



J. M. CLEARY PERCUSSIVE UNIT FOR EARTH DRILLING Sept. 6, was

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PERCUSSIVE UNIT FOR EARTH DRILLING Filed Dec. 6, 1965 5 Sheets-Sheet 4INVENTOR James M. Cleury Mw M Attorney Sept. 6, 1986 J. M. CLEARYPERCUSSIVE UNIT FOR EARTH DRILLING 5 Sheets-Sheet 5 Filed Dec. 8, 1965By James M. Cleory Attorney United States Patent 3,270 822 PERCUSSIVEUNIT FOR EARTH DRILLING James M. Clear-y, Dallas, Tex., assignor to TheAtlantic Refining Company, Philadelphia, Pa., a corporation ofPennsylvania Filed Dec. 6, 1965, Ser. No. 511,870 29 Claims. (Cl.173-137) This is a continuation-in-part of application Serial No.241,412, filed November 30, 1962, now Patent No. 3,229,775.

This invention is concerned with fluid-actuated, percussive units forearth borehole drilling.

Fluid-actuated earth drilling percussive units are commonly composed ofan elongated tubular housing having a power fluid inlet and outlet, 21valving and passage system for controlling passage of the power fluidthrough the power unit, and a power chamber. Slidably mounted in thepower chamber is a piston-like hammer which is adapted to reciprocate upand down to impact an anvil to which a drill bit is attached. Usually,the anvil is splined or grooved to cooperate with grooves or splines inthe lower portion of the housing so that the drill bit and anvil arerotated by rotation of the housing. Such percussive devices are rotatedso that the cutting elements of the drill bit are indexed to a newposition between hammer blows. Longitudinal reciprocal movement of thehammer is accomplished by alternately applying power fluid to the endsof the hammer. When power fluid is applied to the lower end of thehammer between the anvil and hammer, the hammer is accelerated in anupward direction until entry of the power fluid no longer acts on thelower end of the hammer. The upward energy of the hammer is thentransmitted to an elastic rebound cushion. This rebound cushion isusually either a trapped volume of compressible power fluid, or a hammerreturn spring, or both. The hammer continues moving upward until thevelocity of the hammer is zero and the direction of movement of thehammer is reversed. Before or at the peak of the upstroke of the hammer,the valving system switches flow of high pressure power fluid to theupper end of the hammer. When the hammer starts to move downward, partof the energy stored in the elastic rebound cushion is returned to thehammer. This energy combines with energy developed by the power fluid toaccelerate the hammer downward to impact the anvil. This downwardpercussive force is transmitted to the drill bit and .drives the cuttingelements of the drill axially into the earth formation.

The ultimate objective in percussive tool design is to provide a durablegasor liquid-operated tool exhibiting both a high energy, high frequencyimpact and a high fluid conductivity per unit diameter over a wide rangeof operating conditions. The achievement of this ultimate objectivedepends primarily on the stability, reliability and uniformity of thetiming relationship between the up and down movement of the hammer andthe instants that the valving system ceases and starts the flow of powerfluid to alternate ends of the hammer. This timing relationship aflfectsand is affected by almost every operating feature of the percussiveunit; however, it has been found that many of the problems involved ininterrelating all of the tools features to achieve the ultimateobjective are overcome by providing a valving arrangement exhibitingbetter cooperation between the hammer and valve and by providing betterspring means above the hammer for reversing the upward movement of thehammer into its downward or power stroke.

Accordingly, it is an object of this invention to provide a valvingsystem that increases the horsepower output per unit energy input andper unit tool diameter, permits 3,270,822 Patented Sept. 6, 1966 the useof both compressible and incompressible power fluids, increases theconductivity to fluid flow of the power unit, provides more reliable anduniform operation and cooperative relation between the hammer and valve,causes the power fluid to act on the hammer throughout the power stroke,permits the power unit to operate efficiently over a wider range ofoperating conditions, reduces friction and failures, or exhibits greatflexibility of adjustment or timing to diflerent percussive toolcomponents and other operating conditions.

It is a further object of this invention to provide an overhead springfor use during the period when the hammer reverses direction of movementand that allows the valve and hammer to operate more eflicientlytogether while reducing wear of the cooperating parts. This overheadspring provides smoother uniform movement of the hammer and valve,increases timing reliability and stability, causes the full power fluidpressure differential to act on the hammer to the moment of impact,permits greater power unit fluid conductivity, allows the use of bothincompressible and compressible power fluids, increases the operatingrange of the power unit with more efficient horsepower output per unitenergy input per unit tool diameter, or is suitable for prolonged andrapid receipt, storage, and release of energy to and from the hammer.

Yet another object is to provide a system that facilitates removal ofimpurities or other contaminants from the incoming power fluid.

Other advantages and objects of this invention will become apparent byreference to the accompanying drawings, appended claims and followingspecification.

In the drawings:

FIGURES 1 and 2 are elevational, fragmented, partial cross-sectionalviews showing the internal construction of one embodiment of thepercussive mechanism described herein.

FIGURE 3 is an elevational, fragmented cross-sectional view of analternative construction of the embodiment of FIGURE 2.

FIGURE 4 is an elevational, cross-sectional view of the hammer of FIGURE1.

FIGURES 5 and 6 are top and bottom views, respectively, of the hammer ofFIGURE 4.

FIGURE 7 is a side, elevational, cross-sectional view of the hammer ofFIGURE 4.

FIGURE 8 is a fragmented, elevational cross-sectional view of therotatable valve of FIGURE 1.

FIGURE 9 is a fragmented, side, elevational view of the rotatable valveof FIGURE 8.

FIGURE 10 is a top, cross-sectional view taken at 1010 in FIGURE 9showing the symmetry of the exhaust and inlet ports in the valve.

FIGURE 11 is a top, cross-sectional view taken at 11-11 in FIGURE 1 andshowing one embodiment of fluid-actuated rotary drive means for thepercussive mechanism.

FIGURES l2 and 13 are fragmented, elevational, and top cross-sectionalviews of means for removing impurities from the incoming power fluid asshown in FIG- URE 1.

FIGURE 14 is an isometric as shown in FIGURE 1.

FIGURES 15, 16, and 17 are a front, side, and top view, respectively, ofa follower means as shown in FIG- URE 1.

FIGURE 18 is an elevational, fragmented cross-sectional view showing theinternal construction of another embodiment of a percussive mechanismdescribed herein.

FIGURE 19 is an elevational, fragmented cross-sectional view of adouble-acting groove cam inside the hammer of FIGURE 18.

view of a single-acting cam FIGURE 20 is a cam layout for the cam ofFIGURES 18 and 19 FIGURE 21 is a schematic depicting a hammer and valvecycle.

FIGURES 22, 23 and 24 are an elevational cross-sectional, a topcross-sectional and a bottom end view, respectively, of one of the rodsprings of FIGURE 18.

"First, this invention concerns a ported, rotatable tubular valvecombined with motion conversion means adapted to employ thereciprocating or longitudinal movement of a hammer to coordinate theaxial motion and position of the hammer with the circumferentialrotative position of power fluid inlet and exhaust ports in the valveduring at least a portion of the reciprocating cycle of the hammer. Thismotion conversion means may be a cam and follower arrangement. Theoperation of the rotatable valve is vastly improved if the tubular valvehas at least two inlet ports and at least two exhaust ports which arespaced substantially symmetrically around the circumference of thevalve, thereby balancing fluid pressures exerted on the valve. In oneembodiment, the valve is connected to a rotary drive means which rotatesthe valve at a speed of rotation greater than the speed of rotation ofthe percussive unit and drill bit. When the rotary drive means ispresent, the cam may be a singleor double acting cam. In a secondembodiment, the cam is adapted to rotate the valve and a separate rotarydrive means is not required. One means for accomplishing this is to makethe cam. a double-acting cam.

Secondly, this invention concerns a novel elongated, rod-like springmember positioned above a hammer of a percussive unit. The rod-likespring member is at least twelve times as long as its average effectivewidth with its longitudinal axis substantially parallel to thelongitudinal axis of the casing and having a lower end positioned tosquarely or axially contactan upper surface.

of the hammer. The spring receives energy from the hammer during a lastportion of the hammers upstroke and returns a portion of this energyduring a first part of the downstroke or power stroke of the hammer.Preferably, this rod-like spring is combined with the rotatable valveand is restrained against buckling.

Thirdly,

this invention relates to a percussive unit embodying some or all of theabove and other elements of this invention. Other elements of thisinvention not mentioned previously are the employment of a spring abovethe hammer having a spring rate of at least five thousand pounds perinch, and the placement of a contaminant removal means or by-passseparator inside the rotatable tubular valve.

The word rotatable as used herein in relation to the tubular valveincludes a tubular valve that rotates back and forth in an arc of onerevolution or less as well as one that motates in only one direction.

More specifically, in FIGURES 1 and 2, there is shown one embodiment ofa percussive unit built in accordance with the present invention. Thepercussive unit includes tubular housing 11 whose upper end is removablycon-' nected to a rotary drill string (not shown) which conducts powerfluid to the percussive unit.

slidably telescoping into the lower end of housing 11 is anvil 13 havingan upper anvil surface and which is capable of limited longitudinalmovement within the housing. Longitudinally traversing the anvil iscentral anvil bore passage 15 whose lower end (not shown) communicatesin the usual fashion with fluid discharge passages in a drill bit. Inthe usual manner, the anvil is designed to rotate with housing 11.

In housing 11 above anvil 13 is piston-like hammer 17 which is slidablymounted within housing 11 to undergo reciprocal, longitudinal, or up anddown movement and periodically impact anvil 13. The hammer has uppersurface 19 and lower surface 21. In the illustrated embodiment the lowerportion of hammer 17 is of smaller diameter than the upper portion,thereby forming shoulder 4 23 at a point intermediate of the ends of thehammer.

As shown in FIGURES 1 and 4 through 7, longitudinally traversing hammer17 is central bore 25. Communicating with this central bore passage isat least one upper or first hammer passage 27 which also communicateswith the upper surface or end of the hammer and the interior of housing11 above the hammer. As shown, first hammer passages 27 are two verticalgrooves leading to the top of the hammer which are spaced 180 degrees oncenter and substantially symmetrically around the circumference ofcentral bore 25. These passages extend deep enough into the bore of thehammer to communicate with ports in a rotatable valve and, duringoperation of the percussive unit, these first hammer passages conductpower fluid to and from the cylinder chamber above the hammer ashereafter set forth.

Also communicating with the central bore passage in the hammer is atleast one lower or second hammer passage 29 which in turn communicateswith the lower surface or end of the hammer by way of shoulder 23. Asshown, second hammer passages 29 include a groove in the Wall of thecentral bore of the hammer, two bores which pass through the walls ofthe hammer and two.

vertical grooves in the outer surface of the hammer. These passages arespaced 180 degrees on center and substantially symmetrically around thecircumference of central bore 25. These passages are also spaced degreeson center from first hammer passages 27. Second hammer passages 29 arelocated in the hammer bore to communicate with ports in a rotatablevalve and, during operation of the percussive unit, these second hammerpassages conduct power fluid to and from the lower end of the hammer ashereafter set forth. l

Extending longitudinally through the central bore in hammer 17 and intothe central bore of the anvil is a rotatable tubular valve 31 which, asillustrated, has a central passage which is open at its lower end andwhich near its upper end has inlet openings 33. The tubular valve isfree to rotate substantially independently of housing 11 and the anviland bit.

Near the top of the housing and tubular valve is rotary drive means 35which is connected to or made a part of tubular valve 31. The rotarydrive means will be hereafter described in more detail.

Below rotary drive means 35 and inlet openings 33 in tubular valve 31 isat least one first port through the wall of the tubular valve which portor ports form power fluid inlet ports 37 which, when the drill bit is onbottom and the valve is rotated, alternately communicate with the firstand second hammer passages to conduct power fluid to above and below thehammer. Preferably, for reasons herein made apparent, there will be atleast two inlet ports in the tubular valve and the ports will be spacedsubstantially symmetrically around and through the circumference of thetubular valve as shown in FIGURES 1, and 8 through 10, wherein there aretwo such inlet ports which are degrees on center.

As shown in FIGURES 1 and 8 through 10, through the wall of the tubularvalve is at least one second port which port or ports form power fluidexhaust ports 39, which, when the drill bit is on bottom and the valveis rotated, alternately communicate with the first and second hammerpassages to conduct power fluid from above and below the hammer.Preferably, for reasons herein made apparent, there will be at least twosuch exhaust ports and the number of exhaust ports will equal the number:of inlet ports 37. These exhaust ports will be spaced substantiallysymmetrically around and through the circumterence of the tubular valve.As shown, there are two such exhaust ports which are 180 degrees oncenter and are spaced 90 degrees on center from each inlet port.

In tubular valve 31 separating power fluid inlet ports: 37 from fluidexhaust ports 39 is flow restrictive means. 41 which effectively dividesthe central passage through tubular valve 31 into two sections with thesection above the flow restrictive means being a power fluid inletsection and the section below the flow restrictor being a power fluidexhaust section. For illustrative purposes only, the wall of the valveand flow restrictor is equipped with grooves that give greater verticallength to inlet ports 37 and exhaust ports 39.

Flow restrictive means 41 has bleed passage 43 which provides greatercontrol of the conductivity of the percussive unit and permits powerfluid to escape from the power inlet section of the tubular valve to theexhaust section thereby assuring a continuous stream of fluid forcleaning and flushing the drill bit. Bleed passage 43 also provides away to remove impurities from the incoming power fluid as hereafterdescribed.

As stated previously, through the wall of tubular valve 31 above powerfluid inlet ports 37 and below rotary drive means 35 are inlet openings33 which communicate with power fluid passageway 45 in casing 11 whichconduct power fluid from rotary drive means 35 to tubular valve 31.

Circumscribing and supporting tubular valve 31 is hearing 47 whichpermits easy rotation of the valve. Preferably bearing 47 will be aball, thrust bearing.

Return now to rotary drive means 35. This rotary drive may be any formof fluid-actuated rotary device or fluid reaction motor connected to andsuitable for rotating tubular valve 31 and may or may not be an integralpart thereof. As shown in FIGURES l and 11, rotary drive means 35 is ofthe jet type and includes hollow central chamber 49 which communicateswith incoming power fluid from a drill string attached to the upper endof housing 11. Extending outward from central chamber 49 are jetpassages 51 and 53 which terminate in jet openings 55 and 57.

The top of central chamber 49 forms a rotatable seal with insert 59 inhousing 11. Power fluid passing into central chamber 49 passes to jetopenings 55 and 57 located on opposite sides of the rotary drive meansso that fluid emitting from these jet openings will cause the rotarydrive means to rotate.

As described in copending application Serial No. 241,- 412, the rotarydrive means may be equipped with a suitable form of normally closedcentrifugally-operated bypass port which is opened to allow some of thepower fluid to bypass the rotary drive means when the rotary drive meansreaches its desirable operating speed. This makes the tool moreconductive and permits better control of the timing of the valve.

As stated previously, bleed passage 43 is also used as a passageway forremoving contaminants or impurities from the inlet section of thetubular valve. Since tubular valve 31 is rotated almost independently ofthe housing and drill bit, the incoming power fluid spins and theheavier contaminants and impurities in the power fluid are collected onthe inner walls of the inlet section of the rotating tubular valve.These impurities travel down the inner wall of the tubular valve untilthey are collected by a contaminant removal means where the impuritiespass through bleed port 43 bypassing the hammer chamber. As shown inFIGURES 12 and 13, the contaminant removal means is made up of uppercylinder portion 61 and lower cylinder portion 63 through which powerfluid may pass. Lower cylinder portion 63 is of larger diameter thanupper cylinder portion 61 and forms a seal with the inner wall of thetubular valve. This in turn forms trough-like annulus 65 which collectsimpurities traveling down the inner walls of tubular valve 31. Extendingthrough one side of lower cylinder 63 is bleed passage 43 whichcommunicates with trough-like annulus 65 and the exhaust section of thetubular valve below flow restrictive means 41. Impurities collecting inthe troughlike annulus travel into bleed passage 43 where the impuritiesare flushed into the exhaust section of the percussive unit therebybypassing the hammer chamber.

Above hammer 17, housing 11 provide-s a chamber for an elastic reboundcushion which, during a last portion of the upward travel of the hammer,receives and stores energy from the hammer causing the hammer todecelerate and which, during a first part of the downstroke of thehammer, returns a portion of the stored energy to the hammer.

Preferably, the elastic rebound cushion is a compression spring. Animproved coil-type compression spring is shown in FIGURE 1 wherein theimproved coil spring is comprised of first coil spring 67 and secondcoil spring 69 which are threaded together in a double helix so that thecoils of springs 67 and 69 act in parallel vertical alignment, that is,alternate coils of one spring lie just above and below a coil of theother spring .and the coils are aligned. This improved double helix typecoil spring has fewer turns per individual spring and exhibits a springcoefficient greater than a single coil spring that could be fitted inthe available space. The double helix spring is placed in housing 11above hammer 17 and positioned to contact upper surface 19 during a lastportion of the upstroke of the hammer.

An alternate and superior type of compression spring is shown in FIGURES18 and 22 through 24. This superior spring is comprised of at least onecompression rod spring. As shown in FIGURE 18, there are two elongatedcompression bar or rod-like springs 71 and 73 mounted in housing 11above hammer 17. These rod-like springs are positioned and adapted to beloaded axially when struck by upper surface 19 of the hammer;consequently, rod-like springs 71 and 73 have lower ends 75 and 77,respectively, which are positioned above the harnmer to be squarelycontacted and receive energy from the hammer. The sides of each rod-likespring are supported laterally by housing 11 or other support to preventbuckling; consequently, the rod-like springs act only in" axialcompression. This type of compression spring is more uniformly stressedthan a coil spring and operates under much higher fatigue stresses. Thelength, L, of each rod-like spring must be at least twelve times asgreat as the average maximum width, W, of the main body of the rod ifthe rod-like spring is to operate satisfactorily. Much greater length towidth ratios are desired. The spring'coeflicient for the rod-like springwill ordinarily be greater than 10,000 pounds per inch and springcoefficients much greater than this are preferred.

It has been found that the ported, rotatable tubular valve operates muchmore efliciently if it is combined with motion conversion means adaptedto employ the reciprocating or longitudinal movement of the hammer tocoordinate the axial motion and position of the hammer with thecircumferential or rotative position of the ports in the valve during atleast a portion of the reciprocating cycle of the hammer. In otherwords, the motion conversion means converts longitudinal motion intorotary motion. This motion conversion means may be a cam and followerarrangement with the cam having at least one curved surface adapted tocontact the follower during at least a portionof the reciprocating cycleof the hammer. In a cam and follower arrangement, either the cam or thefollower may be in contact with or connected to the tubular valve. Asillustrated in FIGURES 1 and 14, single-acting cam 79 is embedded in theupper end of hammer 17 surrounding tubular valve 31 so that the cammoves up and down with the hammer. The top of singleacting cam 79 hassingle-acting cam surface 81 which has the same general form as therelative motion between the valve and hammer. Since the tubular valveillustrated f has two power fluid inlet ports and two power fluidexhaust ports spaced symmetrically around the valve, there is a hammercycle every degrees of rotation of the valve or two cycles per valverevolution; consequently, there are two single-acting cam surfaces ofthe same shape covering 180 degrees of the cam cylinder.

Above single-acting cam 79 is a follower means which is a member, suchas a pin, roller or second cam surface, adapted to cooperate withlongitudinal movement of single-acting cam 79 during at least a portionof the up and down strokes of the hammer by riding on single-acting camsurface 81 whenever the ports in the tubular valve are not in properposition relative to the up or down position of the hammer. Thus,single-acting cam surface 81 will either speed up or slow down tubularvalve 31 whenever ports 37 and 39 are out of position relative to theposition of the hammer. As shown, the follower means is pin or roller 83which traverses tubular valve 31 extending outward through follower slot85. The follower roller extends beyond the wall of the tubular valve bya distance suflicient to allow the follower roller to be verticallyaligned with single-acting cam surface 81.

Preferably, either the follower or the cam should be shock mounted toallow either one to slip or give way if there is a malfunction whichcauses a miss-match between the up and down motion of the hammer and therotational movement of the tubular valve. This shock mounting helps toprevent failure or breakage of the percussive unit. One method of shockmounting the follower roller 83 is shown in FIGURES 1 and through 17.The follower roller is aflixed to cam follower bushing 87 which in turnis connected to cam follower bushing return spring 89. It will be alsonoted that follower slot 85 is elongated longitudinally along the lengthof tubular valve 31. In this manner, whenever there is a miss-matchbetween cam and follower, follower roller 83 will slip upward infollower slot 85 causing cam follower bushing 87 to compress camfollower bushing return spring 89 which will return the follower to itsoriginal position when the miss-match is corrected or the hammer movesdownward out of contact with the follower.

It will be noted that when a single-acting cam is used, it is usuallynecessary to have a separate means for starting rotation of tubularvalve 31. In the embodiment of FIGURE 1, this separate means is rotarydrive means however, rotary drive of the valve may be accomplished bypositioning the follower to always follow the cam when the tool is inits normal operating position. As illustrated in FIGURES 18 through 20,this is accomplished by double-acting groove cam 91 which is formed inthe wall of the central bore of hammer 17. This groove cam has at leasttwo cooperating curved surfaces. Follower roller 83', traversing tubularvalve 31 in a manner similar to follower roller 83, extends betweenthese curved surfaces. A layout for the double-acting groove cam isshown in FIGURE 20. It will be noted that the lower surface of thegroove cam is similar to single-acting cam surface 81. The upper side orsurface of double-acting groove cam 91 can have any shape so long asthis upper surface properly orients the follower to a starting positionwhenever the bit is off bottom. As shown, the upper surface is similarto the lower surface except that the upper surface has been cut out toallow the follower free movement at points of turn around of the hammer.At the upper point of the groove cam is upper groove 93. Upper groove 93communicates with the double-acting cam groove and extends upward. Thisupper groove allows the hammer to drop whenever the anvil drops belowits normal operating position. In other words, when the bit is offbottom, the follower rides up into upper groove 93 allowing the hammerto drop. As will be explained, this provides a way of stoppingreciprocation of the hammer.

Consider now the operation of the percussive unit of FIGURE 1. First,note that tubular valve 31 is rotated thereby rotating power fluid inletports 37 and power fluid exhaust ports 39. As shown, there are two inletports and two exhaust ports with, the two inlet ports being 180 degreeson center and the two exhaust ports being 180 degrees on center. Alsoeach inlet port is spaced 90 degrees on center from each exhaust port.All of the ports are of substantially the same length and size and, whenrotated, are adapted to communicate with the same cylindrical area.Using the configuration shown, a complete hammer cycle occurs everydegrees of rotation of the valve.

As to the upper and lower passages in the hammer, note that upper orfirst hammer passages 27 and lower or second hammer passages 29reciprocate up and down with hammer 17, as far as the operation of thepercussive unit is concerned, these passages do not rotate when openingand closing to the ports in the rotating valve. A hammer cycle includesone upstroke and one downstroke per every 180 degrees of rotation of thetubular valve.

A diagram representing a typical hammer cycle is shown in FIGURE 21. Inthis diagram, the vertical axis represents vertical movement ordisplacement of the hammer while the horizontal axis represents eithertime or angular movement of the valve. The curve of FIGURE 21,therefore, represents the relative positions of the hammer and valveduring one cycle of the hammer. References to FIGURES l and 21 will aidin an understanding of the operation of the valve and hammer.

It should also be noted that the basic components of the percussive unitare a ported rotating valve for supplying power fluid to and removingpower fluid from the ends of the hammer, a reciprocating hamer withupper and lower passages, a cam and follower, and a rebound spring abovethe hammer.

In the diagram, at time zero, the hammer rests on anvil 13, inlet ports37 are just opening to lower hammer passages 29 and exhaust ports 39 arejust opening to upper hammer passages 27. Power fluid enters rotarydrive drive means 35' via insert 59 and passes outward through jetpassages 51 and 53. The power fluid emits from jet openings 55 and 57causing the rotary drive means to rotate tubular valve 31. Power fluidpasses downward in passageway 45 through inlet openings 33 and into theinlet section of tubular valve 31.. Power fluid inlet ports 37comunicate with lower or second hammer passages 29 which in turncommunicate with the underside surfaces of the hammer. The power fluidpressure acting on the underside surfaces of the hammer causes thehammer to accelerate upward in its up or return stroke. At the sametime, power fluid above the hammer is exhausted through upper firsthammer passages 27, power fluid exhaust ports 39 and the exhaust sectionof tubular valve 31 where the exhausted fluid is conducted through anvilpassage 15 to the drill bit.

Hammer 17 accelerates upward causing single-acting cam surface 81 torise to a point where the cam surface could contact cam follower 83.Whether or not cam follower 83 contacts cam surface 81 will depend onthe angular or rotative position of tubular valve 31. If the valve isrotating either too fast or too slow relative to the position of thehammer, single-acting surface 81 will contact the cam follower causingthe follower to follow the cam surface thereby coordinating the axialmotion and position of the hammer with the angular or circumferentialrotative position of inlet ports 37 and exhaust ports 39 in the tubularvalve.

Hammer 17 continues to accelerate upward in its upstroke as long aspower fluid passes from inlet ports 37 to second hammer passages 29 andthe underside of the hammer. At the desired moment of upward hammertravel and valve rotation, power fluid inlet ports 37 rotate out ofcommunication with second hammer passages 29 cutting off the flow ofpower fluid to the underside of the hammer. In the diagram of FIGURE 21,this point in the hammers cycle occurs when the valve has rotatedangular distance or time X and the hammer has traveled upwardlongitudinal distance B. The exact point in the upward travel of thehammer when the valve shuts off the flow of power fluid will be adjustedto the compressibility of the power fluid, to the type of reboundcushion above the hammer and to other factors.

In the embodiment shown, at the same time as power 9 fluid inlet ports37 rotate out of communication with second hammer passages 29, powerfluid exhaust ports 39 rotate out of communication with first hammerpassages 27; however, it should be realized that the size or shape ofthese ports could be varied to adjust the timing of these occurrences tosuit the power fluid and other operating conditions. For example, if thepower fluid were an incompressible liquid, it would be desirable topermit the fluid above the hammer to escape through power fluid exhaustports 39 until approximately the movement that the hammer reaches itsmaximum upward travel as hereafter described.

When power fluid inlet ports 37 rotate out of communication with thesecond hammer passages, the hammer continues to travel upwarddecelerating and compressing springs 67 and 69 until the upward energyof the hammer is absorbed by these springs and the drill string. Theupward distance traveled by the hammer during this period ofdeceleration is shown in the diagram of FIGURE 21 by the distance C. Thepeak upward distance traveled by the hammer is the distance D.Preferably, the hammer will contact the springs at the same moment asthe flow of power to the underside of ,the hammer is cut off.Consequently, the springs are compressed by the distance C and thesprings remain in contact with the hammer until the hammer travelsdownward a corresponding distance. During this period from the instantthat the hammer contacts the springs to the moment that the hammerleaves the springs, the valve rotates the angular distance or time Y.The importance of Y will be illustrated later.

Before the hammer reaches its upward peak, tubular valve 31 rotates farenough for power fluid inlet ports 37 to communicate with upper or firsthammer passages 27. High pressure power fluid passes upward via passages27 to exert its pressure on the upside surfaces of the hammer. Thispressure causes some deceleration of the hammer and, once the hammerreaches its peak up- Ward travel, causes the hammer to acceleratedownward.

For reasons hereinafter made apparent, it is important that some of theupstroke energy of the hammer be conserved and returned to the hammerwhen the hammer starts to accelerate downward in its power stroke. Theenergy returned to the hammer is stored in returnable or reusable formby the springs above the hammer.

In the embodiment shown, at the same instant as power fluid inlet ports37 rotate into communication with first hammer passages 27, power fluidexhaust ports 39 rotate into communication with lower or second hammerpassages 29. This allows fluid beneath the hammer to exhaust throughsecond hammer passages 29 and power fluid exhaust ports 39 and intoanvil bore 15.

The hammer accelerates downward until it leaves springs 67 and 69. Afterleaving these springs and'during the period Z, the hammer acceleratesdownward until it impacts the anvil. The cycle then repeats itself.

It should be noted that during a major and critical portion of eachhammer cycle when the hammer is changing from its upstroke to itsdownstroke, single-acting cam surface 81 acts as a limit or control onthe position of cam follower 83 to assure that inlet ports 37 andexhaust ports 39 open and close to first and second hammer passages 27and 29 at the optimum moment as determined by the position of thehammer.

The operation of the embodiment illustrated in FIG- URE 18 is similarexcept that the double-acting groove cam 91 causes the valve to rotate,and except for the fact that throughout the cycle, the double-actinggroove acts as a limit or control on the angular position of follower83' relative to the position of the hammer.

There are times when the bit is off-bottom or when the anvil is belowits normal operating position that it is desirable to ceasereciprocation of the hammer while maintaining flow of power fluid. Inthe embodiments of FIGURES l and 18, this is accomplished by placing andproperly spacing annular bypass passage around central bore 25 or 25.During normal operation, annular bypass passage 95 is above inlet ports37 and exhaust ports 39; however, whenever the anvil drops allowing thehammer to also drop below its normal rebound position, annular bypasspassage 95 drops to a position where the bypass passage communicateswith both inlet ports 37 and exhaust ports 39 at the same time. Whenthis happens, the power fluid flows through the valve without causingthe hammer to reciprocate. It should be recalled that, in the embodimentof FIGURE 18, the follower inside double-acting groove cam 91 would keepthe hammer from dropping if it were not for upper groove 93.

In the diagram of FIGURE 21 and the above description of the operationof the percussive unit, a hammer displacement cycle was divided into thethree distinct periods, X, Y, and Z, with the total period per cyclebeing T. During period X, high pressure power fluid acts on the lowersurface areas of the hammer and accelerates the hammer upward a distanceB. At this point, the hammer and valve are properly synchronized by thecam and follower and the pressure exerted by the power fluid is changed.At the start of period Y, the hammer contacts the springs. The hammerloses its upward energy compressing these springs until it reaches thepeak of its upward stroke. The hammer reverses its direction of movementand the springs expand returning energy to the hammer until the hammertravels downward far enough to break contact with the springs. Theperiod Y covers the period that the hammer is in contact with thesprings. During this period Y, the hammer travels a distance of 2C.During the period Z, after the hammer leaves the springs, high pressurepower fluid acting on the upper surface areas of the hammer acceleratesthe hammer downward to impact the anvil.

In a percussive earth drilling system, it is desirable that the hammerimpact energy be close 'as practical to the energy level at which thebit can properly function and that the frequency of these high energylevel impacts be as great as practical. It has been found that thefrequencyand energy level of the impacts is greatly affected by thecharacteristics of the rebound cushion above the hammer and by thetiming relationship between the rotatable tubular valve and hammer. Allof these operating features are closely related to the period Y andaffect most of the objectives originally set forth. They especiallyaflect the uniformity or smoothness of operation of the percussive unitwhich long periods of rapid drilling.

The energy for hammer impact is primarily developed during the periods Xand Z when the power fluid is accelerating the hammer. If some of theupward energy developed during the period X is to be usable as impactenergy, this upward energy must be turned into downward energy. In theembodiment shown, this is accomplished with the springs above the hammerduring the period Y. This process of changing upward energy to downwardenergy many times per minute involves inherent energy losses and theproblems created by cyclic stresses and by rapid repetition ofcompression and expansion. It has been found that these inherent energylosses and associated problems do much more than merely cause acorresponding proportional loss of energy available for use as hammerimpact energy.

For example, it has been found that too low a spring coefl'icient, whichis expressed in force per unit of compression, seriously limits thefrequency of the impacts and causes an excessive loss in impact energy.If the spring coeflicient is too low, the reversal period Y is too long.When the period Y increases, the sum of the periods Y and Z increasesuntil the sum of these periods is greater than X, that is, the timerequired to complete the periods Y and Z is greater than one-half of thehammer cycle. When this occurs, the amount of impact energy lost byincreasing the period Y increases expoare necessary for sustained 1 lnentially. For the embodiment illustrated above, this can best be shownby relating the periods X, Y and Z to the corresponding angular tubularvalve movements. The angular period X is the angle that the valverotates before the flow of high pressure power fluid is switched fromone end of the hammer to the other end, that is, 90 degrees. In otherwords, everytime the valve rotates 90 degrees the flow of high pressurepower fluid is switched. When the hammer takes too much time to reverseits upward movement and the period Y is too long, the sum of Y and Z isgreater than the time that it takes the valve to rotate 90 degrees;consequently, the flow of high pressure power fluid is switched to theunderside of the ham- 7 mer before the downstroke or powerstroke of thehammer is completed. When thevalve switches prematurely, the energyinput during the period Z is decreased and the power fluid acts on theunderside of the hammer counterbalancing downward energy in the hammer.This decelerates the hammer before impact. In fact, if the power fluidwere incompressible or if the sum of Y and Z were excessively large, thehammer would never strike the anvil. Instead the hammer would lose itspotential impact energy to power fluid between the hammer and anvil. Inaddition to causing a large loss of impact energy, premature switchingof the valve causes a lengthening of the reversal time between the powerstroke and upstroke of the hammer in a manner similar to the way thatthe period Y is increased by too low a spring coelficient. Thisincreased reversal time causes adverse effects similar to those causedby increasing the period Y.

For the embodiment illustrated, hammer displacement and the period Y arecontrolled by the springs above the hammer and angular valve movementduring this period by the cam and follower. As shown above, the springscoeificient, which is expressed as force per unit of compression, isimportant to the operation of the percussive unit and the length of theperiod Y. It has been found that this spring coefficient should be atleast as great as 5,000 pounds per inch. The following example will showone reason why this minimum spring coefficient is required.

For the valve illustrated, the period X equals 1 and T equals 2;consequently, the sum of the periods Y and Z should be equal to or lessthan 1. Theoretically, the distance of acceleration up and down for thehammer during periods X and Z is the same and the sum of X and Z equalsthe square root of 2. The period Z, therefore, equals 0.41. The period Yequals 1.0 minus 0.41 or 0.59. As a result, theoretically, the period Yshould be equal to or less than 0.59. however, for compressible powerfluids, the period Y may exceed 0.59 provided that the period Y is notlong enough to cause an excessive decrease in the impact energy andimpact frequency. For example, the period Y could be as great as 0.7 forsome percussive units.

If 0.7 is used as an upper limit for the period Y, a minimum theoreticalspring coeflicient for the spring above the hammer may be calculated.This minimum theoretical coefficient is required to prevent Y fromexceeding the upper 0.7 limit. It should be remembered that, in order tooptimize the impact energy, impact frequency and percussive unitoperation, the spring coeflicient should be much greater than thecalculated minimum because the period Y should be equal to or less thanthe theoretical value 0.59. Also friction and other losses in thepercussive unit tend to increase the period Y. These were not taken intoaccount in determining the minimum theoretical value for Y;consequently, an even greater spring coefliicent is required.

The minimum required spring coefficient may be estimated by thefollowing formula:

where K is the spring coefficient in pounds per inch, P is As apractical matter,

the power fluid pressure differential across the hammer, A is the areaof the hammer acted upon by P, E is the upstroke hammer energy ininch-pounds and Y is the period for turn around of the hammer.

In a typical 4.25-inch O.D. percussive unit, PA is 700 pounds, and E is960 inch-pounds. When Y equals its maximum value of 0.7 and these valuesare substituted in Equation 1, Equation 1 reduces to and K equals 5,000pounds per inch, which as stated previously, is the minimum value for K.

In -a typical 4.25-inch O.D. percussive unit, PA is 700 erally hold truefor other size percussive units because the ratio of piston area to tooldiameter tends to remain a constant and the required impact energy pertool size tends to remain a constant.

Preferably, for reasons noted previously, the spring coeflicient shouldbe much greater than 5,000 pounds per inch. As an example, if a value ofY equals 0.5 is used in Equation 2 instead of Y equals 0.7, the value ofK is 10,000 pounds per inch.

Percussive units for deep earth borehole drilling are such that thereare severe design limitations for the spring above the hammer. There isa limit on available head space above the hammer. There are alsopractical limitations on hammer mass, hammer travel, hammer reversaldistance and time, spring efficiencies, allowable spring stresses,allowable spring energy losses, power fluid pressure differentials,hammer rebound velocity, and the like. These design limitations precludethe use of a stand ard coil spring. For example, when a 4.25-inch O.D.tool was constructed along the lines illustrated in FIGURES 1 and 2 anddesigned to provide suitable impact energy and frequency, the highestcoil spring coeflicient that could be obtained with a single coil springthat exhibited the necessary energy storage capacity together withfatigue resistance was less than 3,500 pounds per inch.

The improved, double helical, parallel-acting coil-type spring shown inFIGURE 1 had a spring coefiicient between 9,000 and 10,000 pounds perinch as compared to a rate of less than 3,500 pounds per inch for asingle coilspring, but even the double parallel coiled spring isdeficient in many ways and is especially deficient when higher springcoefiicients are required to improve the performance of the percussiveunit. In addition to having a high spring coefiicient, the spring shouldexhibit high energy storage per unit mass and volume. The springdisplacement should be closely aligned with the axis of the hammer andthe spring force should be through the center of hammer mass. The springshould also exhibit low fluid dampering characteristics, that is, as thespring compresses and expands, the amount of fluid displacement percycle should be low. Moreover, the spring should display low energylosses due to extraneous vibrations or spring surges. It was discoveredthat an elongated rod-like spring made of materials with highcompressive yield strength that will store energy elastically and placedwith the longitudinal axis of the rod parallel to longitudinal axis ofthe hammer possessed these properties provided that the length of therod was at least twelve times as great as the average maximum width ofthe major length of the bar or rod-like member which does not need to beround. For example, a rod spring having a length of 33.25 inches and awidth or rod diameter of 0.25 inch and a Youngs modulus of 30 millionpounds per square inch had a spring coefficient of 44,400 pounds perinch. This type of compression rod spring is illustrated in FIGURES l8and 22 through 24, and has been described previously. When properlydesigned and positioned, the rod spring operates under much higherallowable stresses. The stress is all compressive and the springmaterial is stressed uniformly. The rod springs exhibit much higherspring coeflicients and storage capacities with much less inherentenergy loss than coil springs. For example, when the hammer reboundsfrom the spring, the spring retains a certain amount of kinetic energy.This retained energy is lost. This lost energy is proportional to thesquare of the hammer velocity and the spring mass. The mass of the rodspring per unit energy is less than the mass of a coil spring requiredto store the same energy. In other words, inherent energy losses in therod spring are much less than inherent energy losses in a coil spring.

From the foregoing discussion of the operation and cycle of the hammerand rotatable tubular valve, it should now be apparent that the timingand smoothness of operation of the valve affects the entire operation ofthe percussive unit including impact energy, impact frequency, wear ofthe operating pants, and the like. One problem encountered with arotating valve is irregular speed of rotation. There are periods whenthe valve tends to rotate either too fast or too slow. Much of thisirregular valve movement is due to friction and the pulsations in powerfluid flow. Some of the friction can be overcome by providing adequateclearance between the tubular valve and hammer, that is, a clearancethat prevents undue rubbing or contact due to play in the bearings;however, there were still intermittent periods of excessive friction andpulsations in the speed of rotation of the valve. It was discovered thatperiods of excessive drag on the rotating valve occur when the powerfluid pressure distribution on the valve is unduly eccentric. Forexample, in a two-ported tubular valve, that is, a valve having only onepower fluid exhaust port and only one power fluid inlet port, the sideof the valve adjacent the power fluid inlet'port is exposed to highpower fluid pressure, while the side adjacent the exhaust port isexposed only to exhaust fluid pressure. A net eccentric force resultswhich is further aggravated by play in the valve bearings. Thiseccentric unbalance has been eliminated by providing greatercircumferential or radial symmetry to the power fluid inlet ports andthe power fluid exhaust ports so as to better equalize the radial forcesdue to different power fluid pressures. This symmetry is accompliShed byproviding two power fluid inlet ports 180 degrees on center and twopower fluid exhaust ports 180 degrees on center and spaced so that eachexhaust port is 90 degrees on center from an inlet port. It would bepossible to use a greater number of inlet and exhaust ports than twoeach. A large number of inlet and extha-ust ports may be especiallyadvantageous in large diameter tools operating at high pressure andfrequencies.

It should be noted further that by increasing the number of ports andplacing them symmetrically around the tubular valve, the frequency andconductivity to fluid flow of the percussive unit is increased. Thisalso reduces the effect of power fluid pressure pulsations on the valve.

While the invention has been described in connection with certainspecific embodiments thereof, it will now be understood that furthermodifications will suggest themselves to those skilled in this art andit is intended to cover such modifications as fall within the scope ofthe foregoing description and appended claims. For example, it should berecognized that the first and second hammer passages or the power fluidinlet and outlet valve ports may be varied in number, location and sizeto adapt the tool to different modifications. One such variation isshown in FIGURE 18 wherein lower or second hammer passages 29' areformed in the inside wall of the central bore through hammer 17'. Asecond such variation is shown in FIGURE 3 wherein the tubular valvedoes not extend into the anvil as it does in FIGURE 2. Instead, theupper end of anvil 13 is equipped with cylinder extension 97 whichextends upward from the anvil int-o control bore 25' of hammer 17". Thelower end of the hammer reciprocates up and down on cylindricalextension 97 which acts as a seal for the lower end of 14 the hammer.This variation would be especially useful for very long liquid-drivenhammers.

I claim:

1. In a power fluid operated percussive unit of a rotary percussivedrilling system having a hammer means and a drilling means and whereinthe percussive unit and the drilling means are rotated and the drillingmeans is impacted against the earth by axial cyclic reciprocation of thehammer means, an improved valving system comprising:

(a) a tubular, rotatable valve having at least one inlet port located insaid valve in a position to alternately pass power fluid to passagescommunicating with the opposite ends of said hammer means as said valveis rotated, and at least one outlet port located in said valve in aposition to alternately pass power fluid from said passagescommunicating with said opposite ends of said hammer means to a powerfluid exhaust passage as said valve is rotated, and

(b) motion conversion means adapted to employ the reciprocating movementof said hammer means to coordinate the axial position of said hammermeans with the circumferential rotative position of said ports in saidvalve during at least a portion of the reciprocating cycle of saidhammer means.

2. The improved valve of claim 1 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around .the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

3. The improved valve of claim 1 wherein the motion conversion means isa cam and follower means, said cam having at least one curved surfacepositioned to contact said follower during at least a portion of thereciprocating cycle of said hammer means.

4. The improved valve of claim 3 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

5. The improved valving system of claim 3 wherein the cam is adouble-acting cam having at least two curved surfaces and the followeris placed between said surfaces of said double-acting cam, saiddouble-acting cam and said follower being adapted to employ thereciprocating movement of said hammer means to rotate said valve uponreciprocation of said hammer means and to coordinate the axial positionof said hammer means with the circumferential position of said ports insaid valve.

6. The improved valve of claim 5 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

7. The improved valving system of claim 1 wherein the rotatable tubularvalve is connected to rotary drive means adapted to rotate said valverelative to the hammer.

8. The improved valving system of claim 7 wherein there are at least twoinlet ports and at least two exhaust ports, said at least two inletports being spaced substantially symmetrically around the circumferenceof said valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

9. The improved valving system of claim 7 wherein the mention conversionmeans is a cam and follower means, said cam having at least one curvedsurface positioned to contact said follower during at least a portion ofthe reciprocating cycle of said hammer.

10. The improved valving system of claim 9 wherein there are at leasttwo inlet ports and at least two exhaust ports, said at least two inletports being spaced substantially symmetrically around the circumferenceof said valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

11. In a percussive earth drilling unit having an elongated tubularcasing, a piston-like hammer having an upper surface and a lowersurface, said hammer being slidably mounted in said casing and adaptedto reciprocate up and down within said casing, an anvil surface slidablymounted in said casing below said hammer and positioned to be impactedby downward movement of said lower surface of said hammer when thepercussive unit is in its normal drilling position, said percussive unitand casing being adapted to receive and discharge power fluid in amanner which causes said hammer to undergo a fluid actuatedreciprocating upstroke and downstroke movement within said casing, aspring mounted within said casing above said upper surface of saidhammer, and positioned to contact said upper surface of said hammerduring a portion of said upstroke of said hammer, said spring adapted toreceive energy from said hammer during the last part of said upstrokeand to return a portion of said energy during the first part of saiddownstroke of said hammer, an improved spring comprising at least oneelongated rod-like member being at least twelve times as long as it iswide with its longitudinal axis substantially parallel to thelongitudinal axis of said casing, said rod-like member having a lowerend positioned to contact said upper surface of said hammer and receiveenergy from said hammer during said last portion of said upstroke and toreturn a portion of said energy during said first part of saiddownstroke.

12. The improved spring of claim 11 wherein the vertical sides of the atleast one rod-like member are supported laterally thereby preventingbuckling of said rodlike member.

13. In rotary percussive earth drilling wherein a per" cussive unit andan earth drill are rotated and the drill is impacted against the earthby axial cyclic reciprocation of a hammer within the percussive unit, apower fluid-operated percussive unit comprising an elongated casing, ananvil surface slidably mounted in the lower end of said casing, apiston-like hammer slidably mounted for reciprocating up and downmovement in said casing above said anvil surface, said hammer having anupper surface and a first hammer passage communicating with said uppersurface, said hammer having a lower surface and a second hammer passagecommunicating with said lower surface, a rotatable tubular valve passinglongitudinally through said hammer, said tubular valve having a valveinlet passage adapted to pass fluid into said valve and a valve outletpassage adapted to pass power fluid from said valve, at least one firstvalve port located in said valve in a position to alternately pass saidpower fluid from said valve inlet passage to said first hammer passageand to said second hammer passage as said valve is rotated, at least onesecond valve port located in said valve in a position to alternatelypass said power fluid from said first and said second hammer passages tosaid valve outlet passage as said valve is rotated, and motionconversion means adapted to employ the reciprocating movement of saidhammer to coordinate the axial position of said hammer with thecircumferential rotative position of said first and said second valveports during at least a portion of the reciprocating cycle of saidhammer.

14. The percussive unit of claim 13 wherein there are at least two firstvalve ports and at least two second valve ports, said at least two firstvalve ports being spaced substantially symmetrically around thecircumference of said valve and said at least two second valve portsbeing spaced substantially symmetrically around the circumference ofsaid valve and from said inlet ports.

15. The percussive unit of claim 13 wherein the motion conversion meansis a cam and follower means, said cam having at least one curved surfacepositioned to contact said follower during at least a portion of thereciprocating cycle of said hammer means.

16. The percussive unit of claim 15 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

17. The percussive unit of claim 15 wherein the cam is a double-actingcam having at least two curved surfaces and the follower is placedbetween said surfaces of said double-acting cam, said double-acting camand said follower being adapted to employ the reciprocating movement ofsaid hammer means to rotate said valve upon reciprocration of saidhammer means and to coordinate the axial position of said hammer meanswith the circumferential position of said ports in said valve.

18. The percussive unit of claim 17 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

19. The percussive unit of claim 13 wherein the rotatable tubular valveis connected to rotary drive means adapted to rotate said valve relativeto the hammer means.

20. The percussive unit of claim 19 wherein the motion conversion meansis a cam and follower means, said cam having at least one curved surfacepositioned to contact said follower during at least a portion of thereciprocating cycle of said hammer.

21. The percussive unit of claim 19 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

22. The percussive unit of claim 21 wherein the motion conversion meansis a cam and follower means, said cam having at least one curved surfacepositioned to contact said follower during at least a portion of therecip rocating cycle of said hammer.

23. The percussive unit of claim 13 wherein there is a spring mountedwith said casing above said upper end of said hammer, said springpositioned to contact. said upper surface of said hammer and receiveenergy from said hammer during a last part of the upward movement ofsaid hammer and to return a portion of said energy during a first partof the downward movement of said hammer, said spring having a springrate of at least 5,000 pounds per inch.

24. The percussive unit of claim 23 wherein the motion conversion meansis a cam and follower means, said cam having at least one curved surfacepositioned to contact said follower during at least a portion of thereciprocating cycle of said hammer means.

25. The percussive unit of claim 24 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said valve andfrom said inlet ports.

26. The percussive unit of claim 13 wherein there is at least oneelongated rod-like spring member mounted within said casing above saidupper end of said hammer,

said spring member being at least twelve times as long as it is widewith its longitudinal axis substantially parallel to the longitudinalaxis of said casing, said rod-like member having a lower end positionedto contact said upper surface of said hammer and receive energy fromsaid hammer during a last part of the upward movement of said hammer andto return a portion of said energy during a first part of the downwardmovement of said hammer.

27. The percussive unit of claim 26 wherein the motion conversion meansis a cam and follower means, said cam having at least one curved surfacepositioned to contact said follower during at least a portion of thereciprocating cycle of said hammer means.

28. The percussive unit of claim 27 wherein there are at least two inletports and at least two exhaust ports, said at least two inlet portsbeing spaced substantially symmetrically around the circumference ofsaid valve and said at least two exhaust ports being spacedsubstantially symmetrically around the circumference of said 20 valveand from said inlet ports.

29. The percussive unit of claim 13 wherein said valve inlet passageinside said tubular valve has a contaminant removal means, saidcontaminant removal means adapted to collect contaminants moving alongthe walls of said valve inlet passage and to discharge said contaminantsdirectly to said valve outlet passage inside said tubular valve.

References Cited by the Examiner UNITED STATES PATENTS 612,969 10/1898Hawkins 9134l 1,333,725 3/1920 Newbert 173137 2,830,293 4/1958 Titus267-10 2,861,519 11/1958 Houle 91341 3,000,197 9/1961 Ruegg et al. 267-13,012,540 12/1961 Vincent et al. 173-64 3,078,827 2/1963 Oelke et al17373 FRED C. MATTERN, 111., Primary Examiner.

L. P. KESSLER, Assistant Examiner.

13. IN ROTARY PERCUSSIVE EARTH DRILLING WHEREIN A PERCUSSIVE UNIT AND ANEARTH DRILL ARE ROTATED AND THE DRILL IS IMPACTED AGAINST THE EARTH BYAXIAL CYCLIC RECIPROCATION OF A HAMMER WITHIN THE PERCUSSIVE UNIT, APOWER FLUID-OPERATED PERCUSSIVE UNIT COMPRISING AN ELONGATED CASING, ANANVIL SURFACE SLIDABLY MOUNTED IN THE LOWER END OF SAID CASING, APISTON-LIKE HAMMER SLIDABLY MOUNTED FOR RECIPROCATING UP AND DOWNMOVEMENT IN SAID CASING ABOVE SAID ANVIL SURFACE, SAID HAMMER HAVING ANUPPER SURFACE AND A FIRST HAMMER PASSAGE COMMUNICATING WITH SAID UPPERSURFACE, SAID HAMMER HAVING A LOWER SURFACE AND A SECOND HAMMER PASSAGECOMMUNICATING WITH SAID LOWER SURFACE, A ROTATABLE TUBULAR VALVE PASSINGLONGITUDINALLY THROUGH SAID HAMMER, SAID TUBULAR VALVE HAVING A VALVEINLET PASSAGE ADAPTED TO PASS FLUID INTO SID VALVE AND A VALVE OUTLETPASSAGE ADAPTED TO PASS POWER FLUID FROM SAID VALVE, AT LEAST ONE FIRSTVALVE PORT LOCATED IN SAID VALVE IN A POSITION TO ALTERNATELY PASS SAIDPOWER